Hydraulic governor having acceleration sensing means



United States Patent [72] Inventor Samuel E. Arnett South Bend, Indiana [21] App1.No. 747,346 [22] Filed July 24, 1968 Patented Sept. 1, 1970 [73] Assignee The Bendix Corporation South Bend, Indiana a corporation of Delaware [54] HYDRAULIC GOVERNOR HAVING ACCELERATION SENSING MEANS 7 Claims, 3 Drawing Figs.

[52] U.S.Cl 123/103, 123/140, 73/512, 137/48 [51] Int. Cl F02d 11/08 Field ofSearch 137/26, 33, 35,48, 54, 56, 58, 34; 123/97,103,103A(1nquired); 73/511, 512, 515 (Inquired) [56] References Cited UNlTED STATES PATENTS 2,407,982 9/1946 Hanna 2,702,560 2/1955 Bobier 137/48 2,754,106 7/1956 lfield 73/512 3,191,610 6/1965 Zeisloff 137/54 3,342,074 9/1967 Stedman..... 73/515 3,450,145 6/1969 Colston 73/515X FOREIGN PATENTS 433,306 8/1935 Great Britain 73/515 Primary Examiner-Clarence R. Gordon Atlorney- Plante, Arens, Hartz and O'Brien COMBUSTION I swzlms 22 o 65 s 67 a I64 45 l 5 54 6 m a P l 6 50 P /74 58 5g 0 I67 p M 54 156 as 2 1 I56 42 52 62 on. I46 I55 60 I28 3 W 144 I45 /04I32 //4 53 P W 55 5 0 46 4 us it I 122 9 2 /34 JZOHG 1 /24 1 a 136 152 3211' W I26 I42 [016, 2 I i PUMP 53 c 4 36 N a0 Patented Sept. 1, 1970 Sheet- T606300 M m 0 M 62 P0 3 3 M w M I A \8. m/\ w oe o M I INVENTOR. SAMUEL E. AENETT AGENT Patented Sept. 1, 1970 Sheet 2 rg QM T Jv z .r R. .T @Q E E2 W v b: 3 av g 5 WA M W M\\\-\\\\ w QN\ E g N W U flillqlll 1| n: W3 Ll mm m m mm Q3 V y k M 0 n9 n2 g m. wwm \Q\ $3 $3 3a m mt. .EP... 1T T N I N@ N W .3 g @2 N w v N5 i 5 um i w 4. m o wt Q3 g .wh 3 .K s 0 R Q 3 mm g M NW NW 3 fi fig .w

AGENT HYDRAULIC GOVERNOR HAVING ACCELERATION SENSING MEANS Various prior art governor devices have been proposed to provide an engine control signal or signals which vary as a function of engine rotational speed and rate of change of engine rotational speed. Most of such prior art devices, for accurate response and/or reliability purposes, utilize somewhat complex arrangements of mechanical elements including centrifugal weights, springs, levers, etc., which are not entirely satisfactory from the standpoint of size, weight and cost as well as maintenance and continued reliability. One example of such prior art governor devices may be found in US. Pat. No. 2,946,188, issued July 26, 1960, to J. M. Eastman (common assignee).

It is an object of the present invention to provide a hydraulic speed governor or relatively simple and compact structure which is capable of providing separate control fluid pressure signals representative of engine speed and rate of change of engine speed.

It is another object of the present invention to provide a hydraulic speed governor wherein a control fluid pressure signal representing engine speed is derived from a pressurized fluid which balances the force exerted by a rotatable centrifugal weight and a separate fluid pressure signal representing rate of change of engine speed is derived from a rotatable coil of fluid filled tubing within which fluid motion occurs as a function of the rotational speed imposed thereon by the engme.

Other objects and advantages of the present invention will be apparent from the following description taken with the accompanying drawings wherein:

H6. 1 is a schematic representation ofa combustion engine and fuel control embodying the present invention therefor;

H6. 2 is a sectional schematic view taken on line 2-2 of F [G 1;

FIG. 3 is a sectional schematic view taken on line 3-3 of FIG. 1.

Referring to FIG. 1, numeral 20 designates a conventional combustion engine having a rotatable output shaft 22 the speed of which represents engine revolutions or speed. Engine speed is controlled in accordance with a control lever 24, the position of which is transmitted via linkage mechanism 26 to a fuel meter generally indicated by 28. A rotary input signal, N, representative of engine rotational speed is supplied to fuel meter 28 via suitable linkage mechanism generally indicated by 30 and connected to output shaft 22. The fuel meter 28, in turn, is connected in flow controlling relationship with a fuel supply conduit 32 which transmits pressurized fuel from an engine driven fuel pump 34 to the engine 20. The inlet of pump 34 is supplied fuel at pressure P from a fuel tank 36.

The fuel meter 28 includes a casing 38 having an inlet 40 connected to receive fuel at pump discharge pressure P, from conduit 32 and an outlet 42 connected to discharge metered fuel flow at pressure P to conduit 32. A variable area fuel pressure responsive metering valve 44 slidably carried in a passage 46 connecting inlet and outlet 40 and 42, respectively, is positioned to vary the effective flow area of passage 46 and thus metered fuel flow at pressure P to the engine 20. The valve 44 is positioned by a differential area piston 48 fixedly secured thereto and slidably carried in a chamber 50. The entire area of one side of piston 48 is exposed to chamber 50 at a controlled servo pressure P,.. The relatively smaller annular area of the opposite side of piston 48 is exposed to chamber 50 which is vented via 2i passage'52 to the interior of casing 38 at relatively low fuel pressure P,,. The interior of casing 38 is vented via passage 53 to the inlet of pump 34 at fuel inlet pressure P Fixed area portions 54 and 55 of metering valve 44 are exposed to fuel pressure P which generates a force aiding the force generated by the fuel pressure P,, acting against piston 48 with the resulting combined forces tending to close valve 44 in opposition to the servo fuel pressure P acting against piston 48 tending to open valve 44. It will be recognized that valve 44 is of the well-known integrating type wherein the valve 44 becomes stabilized when a predetermined ratio of fuel pressures P P and P is established corresponding to the ratio of fixed areas against which the pressure P,, P, and P respectively, act. A variation in one or more of the fuel pressures P P and P will upset the predetermined fuel pressure ratio causing the valve 44 to move in one direction or the other depending upon the relative pressure error and continue to move until the predetermined fuel pressure ratio is established. To that end, the servo pressure P, is derived downstream from a restriction 56 in a passage 58 connecting chamber 50 with passage 46 at fuel pressure l upstream from valve 44. A branch passage 60 connects passage 58 downstream from restriction 56 with the interior of casing 38 at pump inlet pressure P and is provided with a valve orifice 62 at the discharge end thereof. A flapper valve 64 at one end of a lever 66 is adapted to cooperate with orifice 62 to vary the effective flow area thereof and thus the fuel pressure P, intermediate restriction 56 and orifice 62 depending upon the position of lever 66.

The lever 66 is pivotally mounted on a fixed support 67 and is loaded by a reference force derived from a compression spring 68 interposed between lever 66 and a spring retainer 70. The spring retainer 70 is positioned by a follower 72 pivotally mounted on a fixed support 74 and bearing against a rotatable cam 76 which, in turn, is carried by a shaft 78 suitably mounted for rotation in casing 38 and actuated by linkage mechanism 26 in response to movement of control lever 24.

The reference force derived from spring 68 is opposed by a force generated as a function of engine speed, N, by a bellows 80 which has a fixed end secured to a fixed support 82 by any suitable fastening means providing a fluid seal and an opposite movable end 84 bearing against lever 66. The bellows 80 is aligned with spring 68 and vented interiorly via a passage 86 to a source of controlled fuel pressure P which varies with engine speed as will be described. The bellows 80 is exposed exteriorly to fuel pump inlet pressure P, thereby causing the bellows 80 to expand or contract in response to the fuel pressure differential P, -P generated thereacross.

The reference force derived from spring 68 is opposed by a second force generated as a function of the rate of change of engine speed or engine acceleration, N, by a diaphragm 88 pivotally attached to one end of lever 66 via backing plate 89 and pin 90 and exposed on one side to fuel pump inlet pressure P The opposite side of diaphragm 88 is exposed to controlled fuel pressure P} in a chamber 92 partially defined by a cap member 94 which clamps the radially outermost portion of diaphragm 88 to casing 38 via suitable fastening means such as bolts 96.

The casing 38 is provided with a recess 98 and bore 100 adapted to receive a bearing member 102 fixedly secured therein by any suitable means such as a press fit. A rotatable circular table 104 slidably carried by bearing member 102 is provided with a shaft portion 106 which is connected to and driven by linkage mechanism 30 in response to rotary motion of output shaft 22. The table 104 is generally cup-shaped with the base portion thereof bored to provide a diametrically extending cylinder 108 having a fuel inlet port 110 at one end thereof. An annular channel 114 surrounding cylinder 108 Supplies pressurized fuel to inlet port 110 and receives pressurized fuel from passage 46 at pump discharge pressure P via ports 116 and 118 in table 104, an annular recess 120 and passage 122 in bearing member 102 and passage 124 in casing 38. A fixed restriction 126 in passage I24 controls fuel flow therethrough to provide a pressure drop P P,., for control purposes depending upon the effective flow area of inlet port U0. The downstream side of inlet port 110 is beveled to provide a valve seat 128 against which one end of a centrifugal weight member 130 slidably carried in cylinder 108 is adapted to seat thereby controlling the effective flow area of inlet port 110 from which fuel passes to the interior of casing 38 at pressure P via annulus 132 and a plurality of circumferentially spaced apart passages 134 in the wall of cylinder 108 (see FIG. 2), a passage 136 and an annulus 138 in shaft portion 106, a passage 140 in bearing member 102 and passage 142 in casing 38. A plug 144 suitably recessed to accommodate an O ring seal 146 is fixedly secured in one end of cylinder 108 by any suitable means such as a press fit. A stop 148 integral with plug 144 is adapted to be engaged by weight member 130 to limit the axial travel thereof in response to the fuel pressure P acting against the end of weight member 130. The annular channel 114 communicates with the interior of bellows 80 via a passage 150 in table 104, an annulus 152 in bearing member 102, a passage 154 in casing 38 and passage 86.

A double row helically wound tube 156 encircling the wall of cup-shaped table 104 is provided with an open end 158 aligned with the rotational axis of table 104 and supported by an extension 160 of a tubular fitting 162 which, in turn, is axially aligned with table 104 and rotatably carried in a fixed bracket 164 extending from casing 38. The opposite end 165 of tube 156 is secured in an opening 167 in tubularfitting 162 and communicates via passage 166 therein with a passage 168 in bracket 164 which, in turn, communicates through a passage 170 in cap member 94 with chamber 92. A vent passage 172 containing a restriction 174 is provided in bracket 164 to vent passage 168 to the interior of casing 38 at fuel pump inlet pressure P,,.

OPERATION It will be assumed that the engine 20 is stable in operation at a selected speed of output shaft 22 in accordance with the set position of control lever 24. Under such a condition, the lever 66 is held in a torque balanced position as a result of the reference force of compression spring 68 acting through its associated lever arm and the speed generated force derived from bellows 80 acting through its associated lever arm in opposition to spring 68. The diaphragm 88 does not exert a force against lever 66 by virtue of the fuel pressure P being equivalent to the opposing pressure P The servo pressure P, acting against piston 48 is stabilized by the flapper valve 64 which vents fuel from passage 60 as required to maintain the proper fuel pressure ratio between pressures P,, P, and P which stabilizes piston 48. The fuel pressure P transmitted to bellows 80 is generated by the weight member 130 which is urged toward valve seat 128 under the influence of the rotating table 104 to the extent of controlling flow through orifice 110 such that the centrifugal force generated by the rotating weight member 130 as a function of the speed of rotation of table 104 is opposed by an equal and opposite force generated by the fuel pressure P acting against the end area of weight member 130 exposed to annular channel 114.

Now assuming that control lever 24 is actuated to request an engine acceleration to a higher engine speed, the cam 76 is rotated accordingly causing compression of spring 68 and closing movement of flapper valve 64. The resulting increase in servo pressure P, upsets piston 48 and thus metering valve 44 in an opening direction which, in turn, results in an increase in metered fuel flow through outlet 42 and conduit 32 to the engine thereby initiating an increase in speed thereof.

As engine speed increases, the fuel pressure P increases by virtue of the progressive increase in rotational speed of table 104 which tends to urge weight member 130 toward valve seat 128 which, in turn, reduces the effective flow area of orifice 110 causing an increase in pressure P in chamber 114. The pressure P acting against the end of weight member 130 increases as the effective area of orifice 110 decreases with the resulting force progressively balancing the opposing centrifugal force of weight member 130. The bellows 80 being vented interiorly to the pressure P and exteriorly to pressure Pi generates a corresponding force as a result of the P ---P pressure differential acting thereon which force is imposed against lever 66.

As the engine accelerates the rate of change of speed thereof and thus the rate of change of rotation of table 104 and thus tube 156 tends to accelerate the fuel in rotating tube 156 which, due to the tendency of the fuel to resist a change in velocity thereof generates a fuel pressure differential Pi P across tube 156 which varies as a function of the rate of change of rotational speed thereof. The fluid pressure Pf is transmitted via passage 168 to chamber 92 which responds to the Pi -P fuel pressure differential thereacross and loads lever 66 in a clockwise direction in opposition to compression spring 68 thereby augmenting the force applied by bellows 80. It will be recognized that the lead signal represented by pressure differential P P for any given rate of change of speed of output shaft 22 will depend upon the effective diameter of the coil of tube 156 as well as the diameter and length of tube 156 which may be selected in accordance with the control characteristic desired. Furthermore, the length of tube 156 may be minimized by virtue of gain from pressure to torque through the effective lever arm of lever 66 through which diaphragm 88 acts. The cross sectional area of tube 156 should be made sufficiently large to provide the necessary fuel flow into chamber 92 as the volume thereof increases thereby minimizing any tendency for delay of the generated pressure P due to fuel movement in tube 156.

In the event that the rate of change of speed of output shaft 22 tends to exceed a predetermined maximum allowable yalu e, the fuel pressure P and thus fuel pressure differential P will increase accordingly thereby loading lever 66 to the extent that the compression spring 68 is overcome permitting flapper 64 to move in an open direction which, in turn, reduces pressure P, causing a corresponding decrease in metered fuel flow to the engine 20 thereby controlling the rate of change of speed thereof within the predetermined maximum allowable value.

As shown in FIG. 1, the tube 156 is positioned with its one end 158 located at a lower level than the opposite end 165 thereof to generate sufficient pressure head to ensure that the tube 156 is maintained full of fuel,

As the engine approaches the selected higher speed, the combined forces of bellows and diaphragm 88 acting through lever 66 generate a torque which overcomes the torque exerted by compression spring 68 acting throughout its respective lever arm of lever 66 causing flapper 64 to move in an opening direction. The resulting increase in effective flow area of orifice 62 causes a corresponding decrease in pressure P, in chamber 50 thereby reducing movement of metering valve 44 which, in turn, reduces fuel flow to the engine and thus the rate of change of speed thereof. As the acceleration of the engine decreases, the pressure differential P P,, decreases accordingly. Thus, as the engine speed approaches the selected higher speed, the pressure Pi in bellows 80 progressively increases while the pressure Pi, acting against diaphragm 92 progressively decreases thereby ultimately resulting in stabilization of lever 66 and thus flapper valve 64 as well as metering valve 44 controlled thereby in response to a torque balance derived from compression spring 68 and opposing bellows 80.

An engine deceleration initiated by movement of control lever 24 to a position requesting a lower than existing speed results in a sequence which is the reverse of that described above for an engine acceleration. The cam 76 rotates in response to lever 24 thereby relaxing compression spring 68 which, in turn, upsets lever 66 in a direction to open orifice 62 causing metering valve 44 to move in a closing direction threby decreasing fuel flow to the engine. Since the pressure Pi' substantially stabilizes at pressure P, as a result of the constant velocity of the engine at any governed engine speed, a deceleration, for example, from the aforementioned selected higher engine speed to a lower speed results in a corresponding reduction in speed of table 104 and thus tube 156 rotatable therewith. The tendency of the fuel in tube 156 to resist motion thereof through tube 156 results in a corresponding decrease in pressure P,,', which, in turn, generates a pressure differential P,,P across diaphragm 92 which varies depending upon the rate of change of speed of the tube 156. The

As the engine approaches the requested lower speed, the

torque exerted by lever 66 in response to the forces of bellows 80 and diaphragm 92 decreases to the extent that the lever 66 is urged by compression spring 68 in a counterclockwise direction causing a reduction in area of orifice 62 and a corresponding rise in pressure P, in chamber 50. Upon reaching the reggested e n g in e speed, the fuel pressure di ff e regtial P0 P :V across diaphragm 92 has decreased to substantialE zero, thereby eliminatin the anticipation force g lead signal derived from diaphragm 92 as the pressure Py decreases with the result that the lever 66 and thus metering valve 44 becomes stabilized when the opposing torques derived from compression spring 68 and bellows 80 equalize at the requested engine speed.

I claim: 1. Engine governor means for use with a combustion engine fuel control device having engine fuel control valve means responsive to a force error input signal derived from a reference force and an opposing force generated as a function of engine rotational speed, said governor means comprising:

rotatable means operatively connected to the engine for generating said opposing force as a function of engine rotational speed;

means defining a fluid filled passage wound in a helix and operatively connected to said rotatable means for rotation about the axis of said helix; a source of pressurized fluid communicating with at least one end of the two ends of said fluid filled passage; and fluid pressure responsive means operatively connected to said fuel control valve means and the other of said two ends of said fluid filled passage for modifying the response of said fuel control valve means to said force error input signal in response to the fluid pressure generated at said other end which fluid pressure varies as a function of the rate of change of rotational speed of said fluid filled passage and thus fluid contained therein.

2. Engine governor means as claimed in claim 1 wherein:

said other end of said fluid filled passage is vented to said source of pressurized fluid via a restricted passage and said fluid pressure responsive means is responsive to the fluid pressure generated between the restricted portion of said passage and said other end of said fluid filled passage in response to the rate of change of rotational speed of said fluid filled passage. 7

3. Engine governor means as claimed in claim 2 wherein:

said fluid pressure responsive means is provided with opposite sides operatively connected to said restricted passage and responsive to a fluid pressure differential generated thereacross as a result of the rate of change of rotational speed of said fluid filled passage.

4. Engine governor means as claimed in claim 1 wherein:

said rotatable means includes a rotatable casing operatively connected to and rotated by said engine;

a fluid chamber defined by said casing and provided with an inlet and an outlet;

a restricted supply conduit operatively connecting said inlet with a second source of pressurized fluid pressure;

an outlet passage operatively connecting said outlet with said first named source of pressurized fluid which is at a relatively lower pressure compared to said second source of pressurized fluid;

centrifugal weight means slidably carried in said casing for radial movement relative to the axis of rotation of said casing and adapted to control the effective flow area of said outlet and thus the fluid pressure level in said chamber in accordance with he centrifugal force generated by said centrifugal weight means as a function of the rotational speed of said casing;

said weight means being stabilized in response to the fluid pressure in said chamber acting thereagainst in opposition to said centrifugal force; and

fluid pressure responsive means operatively connected to said fuel control valve means and said fluid chamber for generating said opposing force as a function of engine rotational speed.

5, Engine governor means as claimed in claim 4 wherein:

said fluid pressure responsive means communicating with said fluid chamber is a bellows member vented interiorly to said fluid chamber and exteriorly to said first named source of pressurized fluid at relatively lower pressure compared to said second source and responsive to the fluid pressure differential therebetween.

6. Engine governor means as claimed in claim 1 wherein:

said rotatable means includes a rotatably mounted support member; and

said means defining a fluid filled passage is a tube secured to said support member and rotatable therewith about a common axis of rotation.

7. Engine governor means as claimed in claim 4 and further including:

stop means secured to said casing and adapted to be engaged by said centrifugal weight means to limit the radially inward movement thereof in response to the fluid pressure in said chambers. 

